Wspace or process with hot or chilled water are frequently

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1 Related Commercial Resources CHAPTER 12 HYDRONIC HEATING AND COOLING SYSTEM DESIGN Temperature Classifications CLOSED WATER SYSTEMS Method of Design Thermal Components Hydraulic Components Piping Circuits Capacity Control of Load System Low-Temperature Heating Systems Chilled-Water Systems Dual-Temperature Systems Other Design Considerations Other Design Procedures Antifreeze Solutions ATER systems that convey heat to or from a conditioned Wspace or process with hot or chilled water are frequently called hydronic systems. Water flows through piping that connects a boiler, water heater, or chiller to suitable terminal heat transfer units located at the space or process. Water systems can be classified by (1) operating temperature, (2) flow generation, (3) pressurization, (4) piping arrangement, and (5) pumping arrangement. Classified by flow generation, hydronic heating systems may be (1) gravity systems, which use the difference in density between the supply and return water columns of a circuit or system to circulate water; or (2) forced systems, in which a pump, usually driven by an electric motor, maintains flow. Gravity systems are seldom used today and are therefore not discussed in this chapter. See the ASHVE Heating Ventilating Air Conditioning Guide issued before 1957 for information on gravity systems. Water systems can be either once-through or recirculating systems. This chapter describes forced recirculating systems. Successful water system design depends on awareness of the many complex interrelationships between various elements. In a practical sense, no component can be selected without considering its effect on the other elements. For example, design water temperature and flow rates are interrelated, as are the system layout and pump selection. The type and control of heat exchangers used affect the flow rate and pump selection, and the pump selection and distribution affect the controllability. The designer must thus work back and forth between tentative points and their effects until a satisfactory integrated design has been reached. Because of these relationships, rules of thumb usually do not lead to a satisfactory design. Principles Effective and economical water system design is affected by complex relationships between the various system components. The design water temperature, flow rate, piping layout, pump selection, terminal unit selection, and control method are all interrelated. System size and complexity determine the importance of these relationships to the total system operating success. In the United States, present hydronic heating system design practice originated in residential heating applications, where a temperature drop t of 20 F was used to determine flow rate. Besides producing satisfactory operation and economy in small systems, this t enabled simple calculations because 1 gpm conveys 10,000 Btu/h. However, almost universal use of hydronic systems for both heating and cooling of large buildings and building complexes has rendered this simplified approach obsolete. The preparation of this chapter is assigned to TC 6.1, Hydronic and Steam Equipment and Systems. TEMPERATURE CLASSIFICATIONS Water systems can be classified by operating temperature as follows. Low-temperature water (LTW) systems operate within the pressure and temperature limits of the ASME Boiler and Pressure Vessel Code for low-pressure boilers. The maximum allowable working pressure for low-pressure boilers is 160 psig, with a maximum temperature of 250 F. The usual maximum working pressure for boilers for LTW systems is 30 psi, although boilers specifically designed, tested, and stamped for higher pressures are frequently used. Steam-to-water or water-to-water heat exchangers are also used for heating low-temperature water. Low-temperature water systems are used in buildings ranging from small, single dwellings to very large and complex structures. Medium-temperature water (MTW) systems operate between 250 and 350 F, with pressures not exceeding 160 psi. The usual design supply temperature is approximately 250 to 325 F, with a usual pressure rating of 150 psi for boilers and equipment. High-temperature water (HTW) systems operate at temperatures over 350 F and usual pressures of about 300 psi. The maximum design supply water temperature is usually about 400 F, with a pressure rating for boilers and equipment of about 300 psi. The pressure-temperature rating of each component must be checked against the system s design characteristics. Chilled-water (CW) systems for cooling normally operate with a design supply water temperature of 40 to 55 F (usually 44 or 45 F), and at a pressure of up to 120 psi. Antifreeze or brine solutions may be used for applications (usually process applications) that require temperatures below 40 F or for coil freeze protection. Well-water systems can use supply temperatures of 60 F or higher. Dual-temperature water (DTW) systems combine heating and cooling, and circulate hot and/or chilled water through common piping and terminal heat transfer apparatus. These systems operate within the pressure and temperature limits of LTW systems, with usual winter design supply water temperatures of about 100 to 150 F and summer supply water temperatures of 40 to 45 F. Terminal heat transfer units include convectors, cast-iron radiators, baseboard and commercial finned-tube units, fan-coil units, unit heaters, unit ventilators, central station air-handling units, radiant panels, and snow-melting panels. A large storage tank may be included in the system to store energy to use when heat input devices such as the boiler or a solar energy collector are not supplying energy. This chapter covers the principles and procedures for designing and selecting piping and components for low-temperature water, chilled water, and dual-temperature water systems. See Chapter 14 for information on medium- and high-temperature water systems. Copyright 2008, ASHRAE 12.1

2 ASHRAE Handbook HVAC Systems and Equipment Fig. 1 Fig. 1 Fundamental Components of Hydronic System Fundamental Components of Hydronic System CLOSED WATER SYSTEMS Because most hot- and chilled-water systems are closed, this chapter addresses only closed systems. The fundamental difference between a closed and an open water system is the interface of the water with a compressible gas (such as air) or an elastic surface (such as a diaphragm). A closed water system is defined as one with no more than one point of interface with a compressible gas or surface, and that will not create system flow by changes in elevation. This definition is fundamental to understanding the hydraulic dynamics of these systems. Earlier literature referred to a system with an open or vented expansion tank as an open system, but this is actually a closed system; the atmospheric interface of the tank simply establishes the system pressure. An open system, on the other hand, has more than one such interface. For example, a cooling tower system has at least two points of interface: the tower basin and the discharge pipe or nozzles entering the tower. One major difference in hydraulics between open and closed systems is that some hydraulic characteristics of open systems cannot occur in closed systems. For example, in contrast to the hydraulics of an open system, in a closed system (1) flow cannot be motivated by static head differences, (2) pumps do not provide static lift, and (3) the entire piping system is always filled with water. Figure 1 shows the fundamental components of a closed hydronic system. Actual systems generally have additional components such as valves, vents, regulators, etc., but these are not essential to the basic principles underlying the system. These fundamental components are Loads Source Expansion chamber Pump Distribution system Theoretically, a hydronic system could operate with only these five components. The components are subdivided into two groups: thermal and hydraulic. Thermal components consist of the load, source, and expansion chamber. Hydraulic components consist of the distribution system, pump, and expansion chamber. The expansion chamber is the only component that serves both a thermal and a hydraulic function. METHOD OF DESIGN This section outlines general steps a designer may follow to complete system design. The methodology is not a rigid framework, but rather a flexible outline that should be adapted by the designer to suit current needs. The general order as shown is approximately chronological, but it is important to note that succeeding steps often affect preceding steps, so a fundamental reading of this entire chapter is required to fully understand the design process. 1. Determine system and zone loads. Loads are covered in Chapters 27 to 32 of the 2005 ASHRAE Handbook Fundamentals. Several load calculation procedures have been developed, with varying degrees of calculation accuracy. The load determines the flow of the hydronic system, which ultimately affects the system s heat transfer ability and energy performance. Designers should apply the latest computerized calculation methods for optimal system design. Load calculation should also detail the facility s loading profile facility to enhance the hydronic system control strategy. 2. Select comfort heat transfer devices. This often means a coilor water-to-air heat exchanger (terminal). Coil selection and operation has the single largest influence on hydronic system design. Coils implement the design criteria of flow, temperature drop, and control ability. Coil head loss and location affects pipe design and sizing, control devices, and pump selection. For details on coils, see Chapters 22 and Select system distribution style(s). Based on the load and its location, different piping styles may be appropriate for a given design. Styles may be comingled in a successful hydronic system design to optimize building performance. Schematically lay out the system to establish a preliminary design. 4. Size branch piping system. Based on the selection of the coil, its controlling devices, style of installation, and location, branch piping is sized to provide required flow, and head loss is calculated. 5. Calculate distribution piping head loss. Although the criteria for pipe selection in branch and distribution system piping may be similar, understanding the relationship and effect of distribution system head loss is important in establishing that all terminals get the required flow for the required heat transfer. 6. Lay out piping system and size pipes. After preliminary calculations of target friction loss for the pipes, sketch the system. After the piping system is laid out and the calculations of actual design head loss are complete, note the losses on the drawings for the commissioning process. 7. Select pump specialties. Any devices required for operation or measurement are identified, so their head loss can be determined and accounted for in pump selection. 8. Select air management methodology. All hydronic systems entrain air in the circulated fluid. Managing the collection of that air as it leaves the working fluid is essential to management of system pressure and the safe operation of system components. 9. Select pump (hydraulic components). Unless a system is very small (e.g., a residential hot-water heating system), the pump is selected to fit the system. A significant portion of energy use in a hydronic system is transporting the fluid through the distribution system. Proper pump selection limits this energy use, whereas improper selection leads to energy inefficiency and poor distribution and heat transfer. 10. Determine installation details, iterate design. Tuning the design to increase performance and cost effectiveness is an important last step. Documenting installation details is also important, because this communication is necessary for wellbuilt designs and properly operated systems. THERMAL COMPONENTS Loads The load is the device that causes heat to flow out of or into the system to or from the space or process; it is the independent variable to which the remainder of the system must respond. Outward heat flow characterizes a heating system, and inward heat flow characterizes a cooling system. The quantity of heating or cooling is calculated by one of the following means. Sensible Heating or Cooling. The rate of heat entering or leaving an airstream is expressed as follows:

3 Hydronic Heating and Cooling System Design 12.3 q = 60Q a ρ a c p t where q = heat transfer rate to or from air, Btu/h Q a = airflow rate, cfm ρ a = density of air, lb/ft 3 c p = specific heat of air, Btu/lb F t = temperature increase or decrease of air, F For standard air with a density of lb/ft 3 and a specific heat of 0.24 Btu/lb F, Equation (1) becomes The heat exchanger or coil must then transfer this heat from or to the water. The rate of sensible heat transfer to or from the heated or cooled medium in a specific heat exchanger is a function of the heat transfer surface area; the mean temperature difference between the water and the medium; and the overall heat transfer coefficient, which itself is a function of the fluid velocities, properties of the medium, geometry of the heat transfer surfaces, and other factors. The rate of heat transfer may be expressed by q q = 1.1Q a t = UA( LMTD) where q=heat transfer rate through heat exchanger, Btu/h U=overall coefficient of heat transfer, Btu/h ft 2 F A=heat transfer surface area, ft 2 LMTD = logarithmic mean temperature difference, heated or cooled medium to water, F Cooling and Dehumidification. The rate of heat removal from the cooled medium when both sensible cooling and dehumidification are present is expressed by q t = w h where q t = total heat transfer rate from cooled medium, Btu/h w=mass flow rate of cooled medium, lb/h h =enthalpy difference between entering and leaving conditions of cooled medium, Btu/lb Expressed for an air-cooling coil, this equation becomes q t = 60Q a ρ a h which, for standard air with a density of lb/ft 3, reduces to q t = 4.5Q a h Heat Transferred to or from Water. The rate of heat transfer to or from the water is a function of the flow rate, specific heat, and temperature rise or drop of the water as it passes through the heat exchanger. The heat transferred to or from the water is expressed by q w = m c p t where q w = heat transfer rate to or from water, Btu/h m = mass flow rate of water, lb/h c p = specific heat of water, Btu/lb F t =water temperature increase or decrease across unit, F With water systems, it is common to express the flow rate as volumetric flow, in which case Equation (7) becomes (1) (2) (3) (4) (5) (6) (7) where Q w = water flow rate, gpm ρ w = density of water, lb/ft 3 q w = 8.02ρ w c p Q w t For standard conditions in which the density is 62.4 lb/ft 3 and the specific heat is 1 Btu/lb F, Equation (8) becomes q w = 500Q w t Equation (8) or (9) can be used to express the heat transfer across a single load or source device, or any quantity of such devices connected across a piping system. In the design or diagnosis of a system, the load side may be balanced with the source side using these equations. Heat-Carrying Capacity of Piping. Equations (8) and (9) are also used to express the heat-carrying capacity of the piping or distribution system or any portion thereof. The existing temperature differential t, sometimes called the temperature range, is identified; for any flow rate Q w through the piping, q w is called the heatcarrying capacity. Terminal Heating and Cooling Units Many types of terminal units are used in central water systems, and may be classified in several different ways: Natural convection units include cabinet convectors, baseboard, and finned-tube radiation. Older systems may have cast-iron radiators, which are sometimes sought out for architectural restoration. Forced-convection units include unit heaters, unit ventilators, fan-coil units, air-handling units, heating and cooling coils in central station units, and most process heat exchangers. Fan-coil units, unit ventilators, and central station units can be used for heating, ventilating, and cooling. Radiation units include panel systems, unit radiant panels, infloor or wall piping systems, and some older styles of radiators. All transfer some convective heat. These units are generally used in low-temperature water systems, with lower design temperatures. Similarly, chilled panels are also used for sensible cooling and in conjunction with central station air-handling units isolating outdoor air conditioning. Terminal units must be selected for sufficient capacity to match the calculated heating and cooling loads. Manufacturers ratings should be used with reference to actual operating conditions. Ratings are either computer selected, or cataloged by water temperature, temperature drop or rise, entering air temperatures, water velocity, and airflow. Ratings are usually given for standard test conditions with correction factors or curves, and rating tables are given covering a range of operating conditions. Because the choice of terminal units for any particular building or type of system is so wide, the designer must carefully consider the advantages and disadvantages of the various alternatives so the end result is maximum comfort and economy. Most load devices (in which heat is conveyed to or from the water for heating or cooling the space or process) are a water-toair finned-coil heat exchanger or a water-to-water exchanger. The specific configuration is usually used to describe the load device. The most common configurations include the following: Heating load devices Preheat coils in central units Heating coils in central units Zone or central reheat coils Finned-tube radiators Baseboard radiators (8) (9)

4 ASHRAE Handbook HVAC Systems and Equipment Convectors Unit heaters Fan-coil units Water-to-water heat exchangers Radiant heating panels Snow-melting panels Cooling load devices Coils in central units Fan-coil units Induction unit coils Radiant cooling panels Water-to-water heat exchangers Source The source is the point where heat is added to (heating) or removed from (cooling) the system. Ideally, the amount of energy entering or leaving the source equals the amount entering or leaving through the load. Under steady-state conditions, the load energy and source energy are equal and opposite. Also, when properly measured or calculated, temperature differentials and flow rates across the source and loads are all equal. Equations (8) and (9) express the source capacities as well as the load capacities. Any device that can be used to heat or cool water under controlled conditions can be used as a source device. The most common source devices for heating and cooling systems are the following: Heating source devices Hot-water generator or boiler Steam-to-water heat exchanger Water-to-water heat exchanger Solar heating panels Heat recovery or salvage heat device (e.g., water jacket of an internal combustion engine) Exhaust gas heat exchanger Incinerator heat exchanger Heat pump condenser Air-to-water heat exchanger Cooling source devices Electric compression chiller Thermal absorption chiller Heat pump evaporator Air-to-water heat exchanger Water-to-water heat exchanger The two primary considerations in selecting a source device are the design capacity and the part-load capability, sometimes called the turndown ratio. The turndown ratio, expressed in percent of design capacity, is Turndown ratio Minimum capacity = Design capacity (10) The reciprocal of the turndown ratio is sometimes used (for example, a turndown ratio of 25% may also be expressed as a turndown ratio of 4). The turndown ratio has a significant effect on system performance; lack of consideration of the source system s part-load capability has been responsible for many systems that either do not function properly or do so at the expense of excess energy consumption. The turndown ratio has a significant effect on the ultimate equipment and/or system design selection. Note that the turndown ratio for a source is different from that specified for a control valve. Turndown for a valve is comparable to valve rangeability. Whereas rangeability is the relationship of the maximum controllable flow to the minimum controllable flow based on testing, turndown is the relationship of the valve s normal maximum flow to minimum controllable flow. System Temperatures. Design temperatures and temperature ranges are selected by consideration of the performance requirements and the economics of the components. For a cooling system that must maintain 50% rh at 75 F, the dew-point temperature is 55 F, which sets the maximum return water temperature at something near 55 F (60 F maximum); on the other hand, the lowest practical temperature for refrigeration, considering the freezing point and economics, is about 40 F. This temperature spread then sets constraints for a chilled-water system. Pedersen et al. (1998) describe a classic method for calculating the required temperature of chilled water from the psychrometric chart. The entering water temperature follows the relationship t w = 2t ad t wb (11) where t w = Coil entering water temperature t ad = Apparatus dew point t wb = Coil leaving air wet-bulb temperature The designer should note that there are also constraints imposed on the temperature and differential temperature selection by the chiller selection (e.g., refrigerant choice), and on the coil s flow tolerance with respect to heat transfer. Consult with the chiller manufacturer so that performance requirements of the chiller are taken into account. For a heating system, the maximum hot-water temperature is normally established by the ASME Boiler and Pressure Vessel Code as 250 F, and with space temperature requirements of little above 75 F, the actual operating supply temperatures and temperature ranges are set by the design of the load devices. Most economic considerations relating to distribution and pumping systems favor using the maximum possible temperature range t. Expansion Chamber The expansion chamber (also called an expansion or compression tank) serves both a thermal and a hydraulic function. In its thermal function, the tank provides a space into which the noncompressible liquid can expand or from which it can contract as the liquid undergoes volumetric changes with changes in temperature. To allow for this expansion or contraction, the expansion tank provides an interface point between the system fluid and a compressible gas. By definition, a closed system can have only one such interface; thus, a system designed to function as a closed system can have only one expansion chamber. Expansion tanks are of three basic configurations: (1) a closed tank, which contains a captured volume of compressed air and water, with an air/water interface (sometimes called a plain steel tank); (2) an open tank (i.e., a tank open to the atmosphere); and (3) a diaphragm tank, in which a flexible membrane is inserted between the air and the water (a modified version is the bladder tank). Properly installed, a closed or diaphragm tank serves the purpose of system pressurization control with a minimum of exposure to air in the system. Open tanks, commonly used in older systems, tend to introduce air into the system, which can enhance piping corrosion. Open tanks are generally not recommended for application in current designs. Older-style steel compression tanks tend to be larger than diaphragm expansion tanks. In some cases, there may be economic considerations that make one tank preferable over another. These economics usually are relatively straightforward (e.g., initial cost), but there can be significant size differences, which affect placement and required building space and structural support, and these effects should also be considered. Sizing the tank is the primary thermal consideration in incorporating a tank into a system. However, before sizing the tank, air

5 Hydronic Heating and Cooling System Design 12.5 control or elimination must be considered. The amount of air that will be absorbed and can be held in solution with the water is expressed by Henry s equation (Pompei 1981): where x = solubility of air in water (% by volume) p = absolute pressure H = Henry s constant x = p H (12) Henry s constant, however, is constant only for a given temperature (Figure 2). Combining the data of Figure 2 (Himmelblau 1960) with Equation (12) results in the solubility diagram of Figure 3. Fig. 2 Water Henry s Constant Versus Temperature for Air and With that diagram, the solubility can be determined if the temperature and pressure are known. If the water is not saturated with air, it will absorb air at the air/water interface until the point of saturation has been reached. Once absorbed, the air will move throughout the body of water either by mass migration or by molecular diffusion until the water is uniformly saturated. If the air/water solution changes to a state that reduces solubility, excess air will be released as a gas. For example, if the air/water interface is at high pressure, the water will absorb air to its limit of solubility at that point; if at another point in the system the pressure is reduced, some of the dissolved air will be released. In the design of systems with open or plain steel expansion tanks, it is common practice to use the tank as the major air control or release point in the system. Equations for sizing the three common configurations of expansion tanks follow (Coad 1980b): For closed tanks with air/water interface [( v V t V 2 v 1 ) 1] 3α t = s ( P a P 1 ) ( P a P 2 ) (13) Fig. 2 Henry s Constant Versus Temperature for Air and Water (Coad 1980a) Fig. 3 Solubility Versus Temperature and Pressure for Air/Water Solutions Fig. 3 Solubility Versus Temperature and Pressure for Air/Water Solutions (Coad 1980a) For open tanks with air/water interface For diaphragm tanks v V t = 2V 2 s α t v 1 (14) [( v V t V 2 v 1 ) 1] 3α t = s (15) 1 ( P 1 P 2 ) where V t = volume of expansion tank, gal V s = volume of water in system, gal t 1 = lower temperature, F t 2 = higher temperature, F P a = atmospheric pressure, psia P 1 = pressure at lower temperature, psia P 2 = pressure at higher temperature, psia v 1 = specific volume of water at lower temperature, ft 3 /lb v 2 = specific volume of water at higher temperature, ft 3 /lb α = linear coefficient of thermal expansion, in/in F = in/in F for steel = in/in F for copper t =(t 2 t 1 ), F As an example, the lower temperature for a heating system is usually normal ambient temperature at fill conditions (e.g., 50 F) and the higher temperature is the operating supply water temperature for the system. For a chilled-water system, the lower temperature is the design chilled-water supply temperature, and the higher temperature is ambient temperature (e.g., 95 F). For a dualtemperature hot/chilled system, the lower temperature is the chilledwater design supply temperature, and the higher temperature is the heating water design supply temperature. For specific volume and saturation pressure of water at various temperatures, see Table 3 in Chapter 6 of the 2005 ASHRAE Handbook Fundamentals. At the tank connection point, the pressure in closed-tank systems increases as the water temperature increases. Pressures at the expansion tank are generally set by the following parameters: The lower pressure is usually selected to hold a positive pressure at the highest point in the system (usually about 10 psig). The higher pressure is normally set by the maximum pressure allowable at the location of the safety relief valve(s) without opening them.

6 ASHRAE Handbook HVAC Systems and Equipment Other considerations are to ensure that (1) the pressure at no point in the system will ever drop below the saturation pressure at the operating system temperature and (2) all pumps have sufficient net positive suction head (NPSH) available to prevent cavitation. Example 1. Size an expansion tank for a heating water system that will operate at a design temperature range of 180 to 220 F. The minimum pressure at the tank is 10 psig (24.7 psia) and the maximum pressure is 25 psig (39.7 psia). (Atmospheric pressure is 14.7 psia.) The volume of water is 3000 gal. The piping is steel. 1. Calculate the required size for a closed tank with an air/water interface. Solution: For lower temperature t 1, use 40 F. From Table 3 in Chapter 6 of the 2005 ASHRAE Handbook Fundamentals, v 1 ( at 40 F) = ft 3 lb Fig. 4 Example of Manufacturer s Published Pump Curve Using Equation (13), 2. If a diaphragm tank were used in lieu of the plain steel tank, what tank size would be required? Solution: Using Equation (15), v 2 ( at 220 F) = ft 3 lb [( ) 1] 36.5 ( 10 6 )( ) V t = ( ) ( ) = 578 gal V t [( ) 1] 36.5 ( 10 6 )( ) V t = ( ) = 344 gal V t HYDRAULIC COMPONENTS Pump or Pumping System Centrifugal pumps are the most common type in hydronic systems (see Chapter 43). Circulating pumps used in water systems can vary in size from small in-line circulators delivering 5 gpm at 6 or 7 ft head to base-mounted or vertical pumps handling hundreds or thousands of gallons per minute, with pressures limited only by the system characteristics. Pump operating characteristics must be carefully matched to system operating requirements. Pump Curves and Water Temperature for Constant-Speed Systems. Performance characteristics of centrifugal pumps are described by pump curves, which plot flow versus head or pressure, as well as by efficiency and power information, as shown in Figure 4. Large pumps tend to have a series of curves, designated with a numerical size in inches (10 to 13.5 in. in diameter), to represent performance of the pump impeller and outline the envelope of pump operation. Intersecting elliptical lines designate the pump s efficiency. The net positive suction head (NPSH) required line represents the required entering operating pressure for the pump to operate satisfactorily. Diagonal lines represent the required power of the pump motor. The point at which a pump operates is the point at which the pump curve intersects the system curve (Figure 5). In Figure 4, note that each performance curve has a defined end point. Small circulating pumps, which may exhibit a pump curve as shown in Figure 6, may actually extend to the abscissa showing a run-out flow, and may also not show multiple impellers, efficiency, or NPSH. Large pumps do not exhibit the run-out flow characteristic. In a large pump, the area to the right of the curve is an area of unsatisfactory performance, and may represent pump operation in a state of cavitation. It is important that the system curve always intersect the pump curve in operation, and the design must ensure that system operation stays on the pump curve. Fig. 5 + Fig. 6 Fig. 6 Fig. 4 Example of Manufacturer s Published Pump Curve Pump Curve and System Curve Fig. 5 Pump Curve and System Curve Shift of System Curve Caused by Circuit Unbalance Shift of System Curve Caused by Circuit Unbalance Chapter 43 also discusses system and pump curves. A complete piping system follows the same water flow/pressure drop relationships as any component of the system [see Equation (17)]. Thus, the pressure required for any proposed flow rate through the system may be determined and a system curve constructed. A pump may be selected by using the calculated system pressure at the design flow rate as the base point value. Figure 6 illustrates how a shift of the system curve to the right affects system flow rate. This shift can be caused by incorrectly calculating the system pressure drop by using arbitrary safety factors or overstated pressure drop charts. Variable system flow caused by control valve operation, a larger than required control valve, or improperly balanced systems (subcircuits having substantially lower pressure drops than the longest circuit) can also cause a shift to the right.

7 Hydronic Heating and Cooling System Design 12.7 Fig. 7 General Pump Operating Condition Effects Fig. 8 Operating Conditions for Parallel-Pump Installation Fig. 8 Operating Conditions for Parallel-Pump Installation Fig. 9 Operating Conditions for Series Pump Installation Fig. 7 General Pump Operating Condition Effects (Hydraulic Institute) As described in Chapter 43, pumps for closed-loop piping systems should have a flat pressure characteristic and should operate slightly to the left of the peak efficiency point on their curves. This allows the system curve to shift to the right without causing undesirable pump operation, overloading, or reduction in available pressure across circuits with large pressure drops. Many dual-temperature systems are designed so that the chillers are bypassed during winter. The chiller pressure drop, which may be quite high, is thus eliminated from the system pressure drop, and the pump shift to the right may be quite large. For such systems, system curve analysis should be used to check winter operating points. Operating points may be highly variable, depending on (1) load conditions, (2) the types of control valves used, and (3) the piping circuitry and heat transfer elements. In general, the best selection in smaller systems is For design flow rates calculated using pressure drop charts that illustrate actual closed-loop hydronic system piping pressure drops To the left of the maximum efficiency point of the pump curve to allow shifts to the right caused by system circuit unbalance, direct-return circuitry applications, and modulating three-way valve applications A pump with a flat curve to compensate for unbalanced circuitry and to provide a minimum pressure differential increase across two-way control valves As system sizes and corresponding pump sizes increase in size, more care is needed in analysis of the pump selection. The Hydraulic Institute (HI 2000) offers a detailed discussion of pump operation and selection to optimize life cycle costs. The HI guide covers all types of pumping systems, including those that are much more sophisticated than a basic HVAC closed-loop circulating system, and use much more power. There are direct parallels, though. HI s discussion of pump reliability sensitivity includes a chart similar to that shown in Figure 7. HVAC pumps tend to be low-energy devices compared to industrial or process pumps, which might be responsible for a wide variety of different fluids and operating conditions. Select a pump as close as possible to the best efficiency point, to optimize life-cycle costs and maximize operating life with a minimum of maintenance. HI s recommendations are also appropriate for HVAC pump selection. Parallel Pumping. When pumps are applied in parallel, each pump operates at the same head, and provides its share of the system flow at that pressure (Figure 8). Generally, pumps of equal size are used, and the parallel-pump curve is established by doubling the flow of the single-pump curve (with identical pumps). Fig. 9 Operating Conditions for Series-Pump Installation Plotting a system curve across the parallel-pump curve shows the operating points for both single- and parallel-pump operation (Figure 8). Note that single-pump operation does not yield 50% flow. The system curve crosses the single-pump curve considerably to the right of its operating point when both pumps are running. This leads to two important concerns: (1) the pumps must be powered to prevent overloading during single-pump operation, and (2) a single pump can provide standby service of up to 80% of design flow; the actual amount depends on the specific pump curve and system curve. As pumps become larger, or more than two pumps are placed in parallel operation, it is still very important to ensure in the design that the operating system intersects the operating pump curve, and that there are safeties in place to ensure that, should a pump be turned off, the remaining pumps and system curve still intersect one another. Series Pumping. When pumps are operated in series, each pump operates at the same flow rate and provides its share of the total pressure at that flow. A system curve plotted across the series-pump curve shows the operating points for both single- and series-pump operation (Figure 9). Note that the single pump can provide up to 80% flow for standby and at a lower power requirement. Series-pump installations are often used in heating and cooling systems so that both pumps operate during the cooling season to provide maximum flow and head, whereas only a single pump operates during the heating season. Note that both parallel- and seriespump applications require that the actual pump operating points be used to accurately determine the pumping point. Adding artificial safety factor head, using improper pressure drop charts, or incorrectly calculating pressure drops may lead to an unwise selection. Multiple-Pump Systems. Care must be taken in designing systems with multiple pumps to ensure that, if pumps ever operate in either parallel or series, such operation is fully understood and considered by the designer. Pumps performing unexpectedly in

8 ASHRAE Handbook HVAC Systems and Equipment Fig. 10 Compound Pumping (Primary-Secondary Pumping) On the other hand, if the flow capacity of pump 1 is less than that of pump 2, then point A becomes a mixing point because some water must recirculate upward in the common pipe from point B. The temperature of the water entering the load is between the supply water temperature from the source and the return water temperature from the load. For example, if pump 1 circulates 25 gpm of water leaving the source at 200 F, and pump 2 circulates 50 gpm of water leaving the load at 100 F, then the water temperature entering the load is t load = 200 ( 25 50) ( ) = 150 F (16) Fig. 10 Compound Pumping (Primary-Secondary Pumping) series or parallel have caused performance problems in hydronic systems, such as the following: Parallel. With pumps of unequal pressures, one pump may create a pressure across the other pump in excess of its cutoff pressure, causing flow through the second pump to diminish significantly or to cease. This can cause flow problems or pump damage. Series. With pumps of different flow capacities, the pump of greater capacity may overflow the pump of lesser capacity, which could cause damaging cavitation in the smaller pump and could actually cause a pressure drop rather than a pressure rise across that pump. In other circumstances, unexpected series operation can cause excessively high or low pressures that can damage system components. Standby Pump Provision. If total flow standby capacity is required, a properly valved standby pump of equal capacity is installed to operate when the normal pump is inoperable. A single standby may be provided for several similarly sized pumps. Parallel- or series-pump installation can provide up to 80% standby, which is often sufficient. Compound Pumping. In larger systems, compound pumping, also known as primary-secondary pumping, is often used to provide system advantages that would not be available with a single pumping system. Compound pumping is illustrated in Figure 10. In Figure 10, pump 1 can be referred to as the source or primary pump and pump 2 as the load or secondary pump. The short section of pipe between A and B is called the common pipe (also called the decoupling line or neutral bridge) because it is common to both the source and load circuits. In the design of compound systems, the common pipe should be kept as short and as large in diameter as practical to minimize pressure loss between those two points. Care must be taken, however, to ensure adequate length in the common pipe to prevent recirculation from entry or exit turbulence. There should never be a valve or check valve in the common pipe. If these conditions are met and the pressure loss in the common pipe can be assumed to be zero, then neither pump will affect the other. Then, except for the system static pressure at any given point, the circuits can be designed and analyzed and will function dynamically independently of one another. In Figure 10, if pump 1 has the same flow capacity in its circuit as pump 2 has in its circuit, all of the flow entering point A from pump 1 will leave in the branch supplying pump 2, and no water will flow in the common pipe. Under this condition, the water entering the load will be at the same temperature as that leaving the source. If the flow capacity of pump 1 exceeds that of pump 2, some water will flow downward in the common pipe. Under this condition, Tee A is a diverting tee, and Tee B becomes a mixing tee. Again, the temperature of the fluid entering the load is the same as that leaving the source. However, because of the mixing taking place at point B, the temperature of water returning to the source is between the source supply temperature and the load return temperature. Mixing is a primary reason against application of compound pumping systems, particularly in chilled-water systems with constant-speed pumping applied to the primary pump circuit of the source, and variable-speed pumping applied to the secondary system connected to the loads. The issues associated with this are generally source- and control-related. At one time, chiller manufacturers restricted water flow variation through a chiller, despite the fact that designers and system operators wanted this ability, to increase energy operating economy and reduce operating costs by reducing flow. Mixing was an inevitable by-product, reducing the chiller s operating efficiency. This situation was exacerbated when variablespeed drives were applied to pumps, while chiller pumps stayed constant-speed. Eventually, changes to chiller operating controls allowed flow rate through the chiller evaporator to be varied. Some system designs have shifted away from compound pumping to variable-speed, variable-flow primary pumping to enhance system efficiency. Depending on locale, system size, and selection criteria for the chiller, compound pumping may still be beneficial to system design and operating stability. Chiller selections often are limited to maximum velocities of 7 fps, and minimum velocities of 3 fps. Equal-percentage valve characteristics, though, should operate most often at valve strokes less than 80%, which is a flow rate of less than 40%, and less than the 42% that might represent a low-flow limit for the chiller in a 3 fps operation criterion. Variable-speed pumping in both primary and secondary circuits can be applied to reduce mixing that might occur as chillers are sequenced on and off to compensate for the loads. The following are some advantages of compound circuits: They allow different water temperatures and temperature ranges in different elements of the system. Compound pumping can be used to vary coil capacity by controlling coil temperature, which leads to better latent energy control. They decouple the circuits hydraulically, thereby making the control, operation, and analysis of large systems much less complex. Hydraulic decoupling also prevents unwanted series or parallel operation. Large water circuits often experience pressure imbalances, but compound pumping allows a design to be hydraulically organized into separate smaller subsystems, making troubleshooting and operation easier. Variable-Speed Pumping Application Centrifugal pumps may also be operated with variable-frequency drives, which adjust the speed of the electric motor, changing the pump curve. In this application, a pump controller and typically one frequency drive per pump are applied to control the required system variable. The most typical application is to control differential pressure across one or more branches of the piping network. In a common application, the differential pressure is sensed across a control valve or, alternatively, the valve, coil, and branch (Figure 11). The controller is given a set point equal to the pressure loss of the sensed components at design flow. For example, a 5 psi differential pressure was used to size the control valve; the sensor is connected to sense differential pressure across the valve, so a 5 psi set point is given to

9 Hydronic Heating and Cooling System Design 12.9 Fig. 11 Example of Variable-Speed Pump System Schematic Fig. 13 Area) System Curve with System Static Pressure (Control Fig. 11 Example of Variable-Speed Pump System Schematic Fig. 12 Fig. 12 Example of Variable-Speed Pump and System Curves Example of Variable-Speed Pump and System Curves the controller. As the control valve closes in reaction to a control signal (Figure 12) from 1 to 2, the differential pressure across the branch rises as the system curve shifts counterclockwise up the pump curve. The pump controller, sensing an increase in differential, decreases the speed of the pump to about point 3, roughly 88% speed. Note that each system curve is shown with two representations. The system curve is the simple relationship of the flow ratio squared to the head ratio. Under control of a pump controller with a single sensor across the control valve, as shown in Figure 11, the pump decreases in speed until the theoretical zero-flow point, at which the pump goes just fast enough to maintain a differential pressure under no flow. In this case, the speed is about 88%, and the power is about 68%. In operation, expect there to be a difference in performance. Figure 12 shows the change as a step function, opposite of the way in which the pump controller functions. Based on the control method, the controller adjustment settings (gain, integral time, and derivative time), system hydraulics, valve time constant, sensor sensitivity, etc., it is far more likely that a small incremental rise in differential pressure will have a corresponding small decrease in pump speed, as the valve repositions itself in reaction to its control signal. This appears as a small sawtoothed series of system curve and pump curve intersection points, and is not of real importance. What is important is that many factors influence operation, and theoretical variable-speed pumping is different from reality. Fig. 13 System Curve with System Static Pressure (Control Area) Decreasing pump speed is analogous to using a smaller impeller size in the volute of the pump, and the result is that motor power is reduced by the cube of the speed reduction, closely following the affinity laws. (Variable-speed pump curves are shown in Chapter 43.) Design flow conditions should be minimal hours of operation per year; as such, the energy savings potential for a variable-speed pump is great. In the general control valve application, a reasonably selected equal percentage valve with a 50% valve authority should reduce flow by about 30% when the valve is positioned from 100% open (design flow, minimal hours per year) to 90% open. The 10% change in stroke should reduce pump operating power about 70%. Hydronic systems should operate most of the time at flow rates well below design. The coil characteristic suggests that, in sensible applications, a small percentage of flow yields an exceptionally high degree of coil design heat transfer, adequate for most of the year s operation. Depending on system design, direct digital control may also allow more advanced control strategies. There are various reset control strategies (e.g., cascade control) to optimize pump speed and flow performance. Many of these monitor valve position and drive the pump to a level that keeps one valve open while maintaining set point for comfort conditions. Physical operational requirements of the components must be taken into account. There are some concerns over operating pumps and their motor drives at speeds less than 30% of design, particularly about maintaining proper lubrication of the pump mechanical seals and motor bearings. From a practical perspective, 30% speed is a scant 3% of design power, so it may be unnecessary to reduce speed any further. Consult manufacturers for information on device limits, to maintain the system in good operating condition. Exceptional energy reduction potential and advances in variablefrequency drive technology that have reduced drive costs have made the application of variable-speed drives common on closed hydronic distribution systems. Successful application of a variablespeed drive to a pump is not a given, however, and depends on the designer s skill and understanding of system operation. For instance, the designer should understand the system curve with system static pressure, as shown in Figure 13. This control phenomenon of variable-speed, variable-flow pumping systems is often called the control area; Hegberg (2003) describes the analysis used to create these graphs.

10 ASHRAE Handbook HVAC Systems and Equipment The plot represents the system curves of different operation points, created by valve positions, and their intersection points with a pump curve. These discrete points occur under specific imposed points of operation on the control valves of the hydronic system, and represent potential worst-case operation points in the system. Typically, this imposed sequence is that control valves are closed in specific order relative to their location to the controlled pump, and the total dynamic head of the pump is calculated at each position. When there is one sensor of control, the boundaries as shown are created. The upper system boundary represents system head as valves are closed sequentially, starting with those nearest the pump and ending with those closer to the sensor. The lower boundary curve represents system head and flow when valves are closed sequentially from the location of the sensor toward the controlled pump. The importance of these two boundaries is that, for any given system flow, there are an infinite number of system heads that may be required, based on which control valves are open and to what position. This is opposite of the system curve concept, which is one flow, one head loss in a piping system. The graph is a representation of a controlled operation; the piping system follows the principles of the Darcy equation, but the intervening feedback control adjustments to the pump speed produce an installed characteristic in the system different from what might otherwise be expected. The implication of any boundary above the generic system curve relationship to design flow is that, as flow decreases in the system, it does not follow the affinity law relationship in reducing operating power. Conversely, the boundary below the system curve operates with a greater reduction in power than the generic system curve. Although operating below the system curve may seem to be advantageous, saving energy and pumping costs, the designer is cautioned to remember that these represent a system flow less than that required for comfort control. This operating series of conditions represents control valve interaction. One valve flow affects flow to all other circuits between it and the pump. This interaction can be quite large. Depending on distribution system head loss and system balancing techniques, one valve may reduce system flow by two to three times the individual valve s rated flow. This can negatively affect the control and sequencing of the source, and possibly also the control of the affected terminal units. Controlling this interaction is done by engineering the hydraulic losses and distribution pipe friction head losses, using pressure-independent valves (both balancing and control), using sensors at critical points of the hydronic system, and combinations of these strategies. These issues are part of what has caused debate over system balancing, pump method of control, etc. Simply put, system design dictates the requirements of adjustment and control. Balance of a system is more than having a balancing valve; it involves pipe selection and location combined with valves and coils and, in some cases, pumps, as well as measurement and adjustment devices. Similarly, the effects of one method of pump control over another (e.g., differential pressure control, valve position of pump speed reset) must also be considered. Traditional system design and feedback control techniques use a series of single-loop controllers. These techniques also work on variable-speed pumping systems, but the interaction of all of the individual controlled systems requires analysis by the designer and calculated design choices based on a thorough understanding of the loads, both theoretical and practical, and should also take into account that operators may run systems in a manner not intended by the designer. Address these issues during design, because they are difficult to understand in the field, and the required instrumentation to monitor the effects is rarely installed. Despite these challenges, application of variable-speed, variable-flow pumping can be very satisfactory. However, variable-speed pumping designs require that the designer allow adequate time in calculation and design iteration to mitigate potential operation issues, and ensure the design criteria of a comfort providing system with operating energy and cost efficiency. Pump Connection Pump suction piping should be at least as large as the nozzle serving the pump, and there should be minimal fittings or devices in the suction to obstruct flow. Typically, pump manufacturers prefer five to eight pipe diameters of unobstructed (i.e., no fittings) straight pipe entering the pump. Fittings such as tees and elbows, especially when there is a change in planar direction, cause water to swirl in the pipe; this can be detrimental to pump performance, and may also lead to pump damage. Manufacturers may recommend special fittings in applications where piping space is unavailable to overcome geometry factors. These fittings need to be carefully reviewed in application. Piping to the pump should be independently supported, adding no load to the pump flanges, which, unless specifically designed for the purpose, are incapable of supporting the system piping. Supporting the pipe weight on the flange can cause serious damage (e.g., breaking the flange), or may induce stresses that misalign the pump. Similar results can also occur when the pipe and pump are improperly aligned to each other, or pipe expansion and contraction are unaccounted for. Flexible couplings are one way to overcome some of these issues, when both the pump and the pipe are supported independently of each other, and the flexible coupling is not arbitrarily connected to the pump and pipe. When pumps are piped in parallel, these requirements are extended. Pump entering and discharge pressures should be equal in operation. In addition to maintaining the recommended straight inlet pipe to the pump, manifold pipe serving the inlets should also have a minimum of two manifold pipe diameters between pump suction center lines. Pump discharge manifolds should be constructed to keep discharge velocities less than 10 to 15 fps, and lower when a check valve is applied, because it is necessary to prevent hydraulic shock (water hammer). Soft-seating discharge check valves are required on the pumps to prevent reverse flow from one pump to another. Various specialty valves and fittings are available for serving one or more of these functions. Distribution System The distribution system is the piping connecting the various other components of the system. The primary considerations in designing this system are (1) sizing the piping to handle the heating or cooling capacity required and (2) arranging the piping to ensure flow in the quantities required at design conditions and at all other loads. The flow requirement of the pipe is determined by Equation (8) or (9). After t is established based on the thermal requirements, either of these equations (as applicable) can be used to determine the flow rate. First-cost economics and energy consumption make it advisable to design for the greatest practical t because the flow rate is inversely proportional to t; that is, if t doubles, the flow rate is reduced by half. The three related variables in sizing the pipe are flow rate, pipe size, and pressure drop. The primary consideration in selecting a design pressure drop is the relationship between the economics of first cost and energy costs. Once the distribution system is designed, the pressure loss at design flow is calculated by the methods discussed in Chapter 36 of the 2005 ASHRAE Handbook Fundamentals. The relationship between flow rate and pressure loss can be expressed by Q = C v p (17) where Q=system flow rate, gpm p =pressure drop in system, psi C v = system constant (sometimes called valve coefficient, discussed in Chapter 42) Equation (17) may be modified as follows: